Method for calibrating a detector in a damping assembly

ABSTRACT

A method for calibrating a detector of a damping assembly in a vehicle is disclosed. The damping assembly comprises a damper for connecting a transmission shaft of a transmission system of the vehicle to a driving shaft, the method comprising the step of applying different torques at the damping assembly by generating losses at different working conditions of the transmission system, so that the damper can be compared to a damper of a reference vehicle for each of said different working conditions.

FIELD OF THE INVENTION

The invention relates to a method for calibrating a detector included in a damping assembly connecting two rotatable shafts in a vehicle, particularly a working vehicle such as a tractor, an excavator, a crawler vehicle or the like.

BACKGROUND OF THE INVENTION

Known tractors comprise an internal combustion engine for generating torque in order to rotate a driving axle and possibly drive one or more external devices attached to a power take-off of the tractor. A transmission system is provided for transmitting power from the engine to the driving axle and/or to the power take-off. The transmission system may comprise a continuously variable transmission (CVT), particularly of a hydrostatic type.

A crankshaft of the engine is connected to a shaft of the transmission system by a damping assembly. The damping assembly comprises a damper that is used to deflect and absorb engine vibrations, in order that torque may be transmitted to the transmission system more smoothly.

Angular speed of the crankshaft varies according to a periodic waveform. Angular speed of the transmission shaft downstream of the damping assembly also varies according to a periodic waveform. However, due to the damping assembly, a phase difference exists between angular speed of the crankshaft and angular speed of the transmission shaft.

The damping assembly may comprise a detector for detecting torque which is transmitted from the engine to the transmission system. The detector measures the phase angle at the damping assembly, i.e. the phase difference between angular speed of the crankshaft and angular speed of the transmission shaft.

Once the phase angle has been measured, torque at the damping assembly can be determined by means of a characteristic curve which, for a given damper, gives the relationship between phase angle and torque.

The detector of the damping assembly needs to be calibrated, in order to measure the real value of the phase angle and hence obtain the real value of the torque, in any working condition. The value of torque so obtained can be subsequently used for a number of different purposes.

A calibration method is known in which the phase angle is measured at two different working points, with the clutch completely or almost completely engaged and the brakes activated. The two working points at which the phase angle is measured correspond to different values of torque generated by the engine. High torque values are selected for the two working points, such as, for example, 300 N·m and up to 1000 N·m respectively. The transmission system is therefore subjected to considerable stress.

Furthermore, since the known calibration method is carried out while the clutch is engaged, wear on the clutch components is produced.

Finally, the known calibration method is not particularly accurate, both because only two working points are taken into consideration and because of the manner in which data relating to the two working points are processed.

An object of the invention is to improve known vehicles, particularly working vehicles such as wheeled or crawler tractors and excavators.

Another object is to provide a method for calibrating a detector included in a damping assembly of a vehicle, in which the components of the vehicle and particularly the transmission system thereof are subjected to reduced stress.

A further object is to provide a method for calibrating a detector included in a damping assembly of a vehicle, which involves substantially no wear on the clutch.

Another object is to provide a method for calibrating a detector included in a damping assembly of a vehicle, which produces accurate and reliable results.

SUMMARY OF THE INVENTION

According to the invention, there is provided a method for calibrating a detector of a damping assembly in a vehicle, the damping assembly comprising a damper for connecting a transmission shaft of a transmission system of the vehicle to a driving shaft, characterized in that the method comprises the step of applying different torques at the damping assembly by generating losses at different working conditions of the transmission system, so that said damper can be compared to a damper of a reference vehicle for each of said different working conditions.

The losses so generated correspond to different loads applied to the side of the damping assembly facing the transmission system. By varying the losses in the transmission system, it is possible to define a plurality of working conditions in which the damper associated to the detector to be calibrated can be compared to a damper of a reference vehicle. Each working condition corresponds to a calibration point for calibrating the detector. The latter can thus be calibrated in any desired number of calibration points, thereby increasing accuracy of the method according to the invention.

Since different torques at the damping assembly are applied by generating losses in different working conditions of the transmission system, there is no need to have the transmission system connected to an external load. The method can therefore be carried out on a vehicle whose clutch is open, i.e. totally disengaged.

As a consequence, the method according to the invention does not produce high stresses on the transmission system, nor does it cause wear on the clutch.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention will be better understood and carried out with reference to the attached drawings, which show some exemplificative and non-limitative embodiments thereof, in which:

FIG. 1 is a schematic drawing showing a transmission system of a vehicle;

FIG. 2 is the first part of a block diagram showing the steps of a method for testing a hydrostatic transmission of the transmission system of FIG. 1;

FIG. 3 is the second part of the block diagram of FIG. 2;

FIG. 4 is a graph showing how torque and clutch current vary during the steps of the method shown in FIGS. 2 and 3;

FIG. 5 is a graph showing schematically the variation of a parameter indicative of the volumetric efficiency of the hydrostatic transmission;

FIG. 6 is a phase angle-torque characteristic of a damper, for a reference damper (solid lines) and a damper which is being tested (dotted lines);

FIG. 7 is a schematic drawing showing a transmission system of a vehicle, in an alternative embodiment.

DETAILED DESCRIPTION OF THE INVENTION

FIG. 1 shows a transmission system 1 for transmitting power between an engine 2 and a driving axle of a vehicle.

The vehicle in which the transmission system 1 is inserted can be a working vehicle, such as a wheeled tractor, a crawler vehicle, an excavator or the like.

In the embodiment shown in FIG. 1, the vehicle which incorporates the transmission system 1 is a wheeled tractor having wheels mounted on the driving axle. These wheels have been schematically represented as a rectangle in FIG. 1 and are designated by the reference numeral 3.

It is however intended that what will be explained herebelow is applicable also to crawler vehicles, in which case the driving axle will move a pair of driving sprockets of respective tracks.

The engine 2 can be an internal combustion engine, particularly a diesel engine.

The transmission system 1 comprises a continuously variable transmission (CVT) interposed between the engine 2 and the wheels 3. The continuously variable transmission comprises a hydrostatic transmission including a hydrostatic unit 5, whose outline has been shown schematically by a dashed line in FIG. 1. The hydrostatic unit 5 in turn comprises a hydraulic pump 6 and a hydraulic motor 7 so configured as to be driven by the hydraulic pump 6.

The hydraulic pump 6 can be a variable displacement pump. In particular, the hydraulic pump 6 can be an axial pump and can comprise a swash plate cooperating with a plurality of axial pistons.

An adjusting device 8 is provided for adjusting the position of the swash plate, i.e. for adjusting the swivel angle of the swash plate and consequently the displacement volume of the hydraulic pump 6. The adjusting device 8 can comprise, for example, an electro-valve.

The hydraulic pump 6 and the hydraulic motor 7 are connected to one another by means of a first line 9 and a second line 10. A hydraulic fluid can be sent from the hydraulic pump 6 to the hydraulic motor 7 through the first line 9. In this case, the hydraulic fluid comes back from the hydraulic motor 7 to the hydraulic pump 6 through the second line 10. The first line 9 is therefore a high-pressure line, whereas the second line 10 is a low-pressure line, because the pressure of the hydraulic fluid in the first line 9 is higher than the pressure of the hydraulic fluid in the second line 10.

If however the rotation direction of a shaft of the hydraulic pump 6 is inverted, while all the other working conditions remain unchanged, the hydraulic fluid can also be sent from the hydraulic pump 6 to the hydraulic motor 7 through the second line 10, and come back to the hydraulic pump 6 through the first line 9. The first line 9 is in this case a low-pressure line, whereas the second line 10 is a high-pressure line.

A damping assembly 31 is provided for connecting an engine shaft 4 of the engine 2 to a transmission shaft or main shaft 12 of the transmission system 1. The engine shaft 4 acts as a driving shaft since it rotatingly drives the main shaft 12 through the damping assembly 31. The engine shaft 4 can be a crankshaft of the engine 2.

The damping assembly 31 serves for deflecting and thus absorbing the power pulses generated by the engine 2, so that torque delivered to the main shaft 12 is more constant over an engine cycle.

The damping assembly 31 may comprise a first rotatable element connected to the engine shaft 4 and a second rotatable element connected to the main shaft 12. The second rotatable element can be a damper 13, for example comprising a plurality of resilient elements acting in a circumferential direction to exert a damping action.

The first rotatable element could be, for example, a flywheel.

More detailed information concerning the structure of the damper 13 can be found in EP 0741286, which relates to a mechanical damper. In the alternative, other kinds of damper could be used, for example a viscous damper.

The hydraulic pump 6 has an input shaft 11 that is mechanically connected to the engine 2, so that the input shaft 11 can be rotatingly driven by the engine 2.

If the input shaft 11 is rotated while the swash plate of the hydraulic pump 6 is in a swivelled configuration, energy is transmitted from the hydraulic pump 6 to the hydraulic motor 7.

The input shaft 11 can be connected to the engine 2 via a mechanical connection comprising, for example, a toothed wheel 14 fixed relative to the input shaft 11.

The toothed wheel 14 engages with a further toothed wheel 15 which is fixed relative to the main shaft 12.

An intermediate gear 16 can be interposed between the toothed wheel 14 and the further toothed wheel 15.

The hydraulic motor 7 has an output shaft 17 suitable for being rotated when the hydraulic fluid is sent into the hydraulic motor 7.

A sensor 18 is provided for measuring the speed of the output shaft 17, particularly the rotational speed thereof. The sensor 18 can be associated to a cogwheel 19 fixed relative to the output shaft 17, so that the sensor 18 is adapted to measure the rotational speed of the cogwheel 19 and hence of the output shaft 17.

A detector 20 is provided for detecting one or more working parameters of the main shaft 12, particularly at the damper 13. The detector 20 can be configured for detecting torque transmitted to the main shaft 12 by the damper 13.

A control device which is not shown is provided for controlling the engine 2. The control device is capable of monitoring several working parameters of the engine 2, for example torque and rotational speed of the engine shaft 4.

A clutch 21 allows the wheels 3 to be selectively connected to, or disconnected from, the engine 2.

A mechanical transmission device is interposed between the engine 2 and the clutch 21. In the embodiment shown in FIG. 1, the mechanical transmission device comprises a planetary gearing 22.

The planetary gearing 22 comprises an annulus or outer ring 23 capable of being rotated by the output shaft 17 of the hydrostatic transmission. To this end, the outer ring 23 may be fixed relative to an intermediate toothed wheel 24 arranged to engage with the cogwheel 19.

The planetary gearing 22 further comprises a plurality of planet gears 25 supported by a planet carrier 26.

The planetary gearing 22 further comprises a sun gear 27 which can be fixed relative to the main shaft 12. A further sun gear 28 is also provided, which can engage with the planet gears 25.

Power can be transmitted to the wheels 3 alternatively via the planet carrier 26 or via the further sun gear 28, in which case the planet carrier 26 is left free to rotate.

A synchronizing device 29 is interposed between the planetary gearing 22 and the clutch 21 for allowing a smooth engagement of the gears of the transmission system 1.

More than one synchronizing device can be present, although they have not been shown since they are not relevant for performing the method that will be disclosed below. For example, a further synchronizing device that is not shown can be associated to a tubular element 30 fixed relative to the further sun gear 28.

Torque generated by the engine 2 is split into two torque fractions which reach the planet gears 25 through two different input torque paths. A first input torque path goes from the engine 2 to the main shaft 12 via the engine shaft 4, then to the hydrostatic unit 5 through the toothed wheels 14, 15 and finally to the outer ring 23 of the planetary gearing 22 through the cogwheel 19 and the intermediate toothed wheel 24. A second input torque path goes from the engine shaft 4 to the planet carrier 26 through the main shaft 12 and the sun gear 27. The two input torque paths join to one another at the planet gears 25.

Torque exits from the planetary gear through two alternative output torque paths. The first output torque path passes through the planet carrier 26, the tubular element 30 and the synchronizing device 29. The second output torque path goes from the further sun gear 28 to the further synchronizing device that is not shown.

In an alternative embodiment, only one torque path can be provided for transmitting torque through the planetary gearing 22.

FIGS. 2 and 3 show the steps of a testing method for testing the hydrostatic transmission while the hydrostatic unit 5 is still installed on the vehicle. The testing procedure can be used particularly if a fault in the hydrostatic unit 5 is suspected, in order to check whether the hydrostatic unit 5 needs to be replaced or not. More in general, by means of the testing procedure it is possible to evaluate how the hydrostatic unit 5 is working, i.e. which is the quality of the hydrostatic unit 5.

The vehicle to be tested is kept still. To this end, it is possible to activate a hand brake of the vehicle, so that any movement of the vehicle is prevented.

For safety reasons, the vehicle is resting on a flat surface.

First of all, there is provided a setting step in which the vehicle is set in a test configuration by setting certain working parameters to a predefined value. The setting step is indicated by the reference numeral 40 in FIG. 2.

In the setting step 40, the engine 2 is started and its rotational speed is set to a determined test value. This test value should not be too low in order to have sufficient torque, nor should it be too high in order to avoid wasting power. For example, for the vehicles that have been studied, the test value of the rotational speed of engine 2 could be between 1000 and 1500 rpm.

Furthermore, a desired value of the swivel angle of the swash plate in the hydraulic pump 6 is selected. The selected value of the swivel angle can be expressed by the parameter α, which is equal to the ratio between the current value of the swivel angle and the maximum possible value of the swivel angle. The parameter α can vary between −1 and +1.

In the embodiment that will be described below, the parameter α has been set to +1. In other words, the swash plate of the hydraulic pump 6 has been positioned in its most tilted configuration. This is done by sending a high current to the adjusting device 8, particularly to the corresponding electro-valve, so that the swash plate is swivelled in its most tilted configuration.

It has been experimentally found that, in this way, the swash plate can be kept in a stable position for the whole duration of the testing procedure, by simply saturating current sent to the electro-valve of the adjusting device 8 until the testing procedure is completed.

In principle, however, it is also possible to use different values of α, particularly α=−1.

By setting α to a desired value, the ratio between the output speed and the input speed of the hydrostatic unit 5, i.e. the ratio between the rotational speed of the output shaft 17 and of the input shaft 11 is univocally determined, for a given hydrostatic unit and for given pressure conditions.

In the setting step 40, the synchronizing device 29 is also engaged and set in a desired position, depending on the kind of vehicle to be tested. For example, the synchronizing device 29 can be so positioned as to select a first forward gear.

The clutch 21 is open, i.e. completely disengaged.

A calibration procedure can also be carried out at this point, in order to calibrate the damper 13. The calibration procedure will be described in detail later.

After the setting step 40, there is provided a check step 41 in order to check whether the vehicle has reached stable conditions. In the affirmative, the testing procedure can be continued. Otherwise, it is necessary to wait until stable conditions are reached.

The operations that have been described above define an initial part of the testing method that can be considered as a preliminary transient part in which substantially no load is applied to the transmission system 1. This initial part of the testing procedure is indicated by the letter A in FIG. 4.

FIG. 4 shows with a continuous line how the torque M generated by the engine 2 and measured by the detector 20 varies over time t. In the initial part A of the testing procedure, the torque M varies more or less randomly due to transient phenomena that occur in this phase.

When stable conditions are reached, a first measuring step 42 can take place, as will be explained in detail herebelow.

At this time, the clutch 21 is completely disengaged and the wheels are stationary. No load is applied to the transmission system 1. Therefore, the pressure difference Δp between pressure of the hydraulic fluid in the high-pressure line 9 (or 10) and pressure of the hydraulic fluid in the low-pressure line 10 (or 9) is very low.

In the first measuring step 42, the rotational speed n_(HST0) of the output shaft 17 of the hydrostatic transmission is measured by means of the sensor 18. The rotational speed n_(E0) of the engine shaft 4 is also recorded by means of the control device which controls the engine 2 and hence monitors the rotational speed thereof.

It is possible to record several values of n_(HST0) and of n_(E0) over a period of time, so that they can subsequently be averaged and processed.

The first measuring step 42 occurs during the period indicated by the letter B in FIG. 4. In this period, the torque M generated by the engine 2 is substantially constant. The period B can therefore be considered as a stable phase in which no load is applied to the transmission system 1.

After the first measuring step 42, the testing method comprises an intermediate step 43 in which an interaction with the clutch 21 is provided as will be set out below.

In the intermediate step 43, the clutch 21 is partially engaged, for example by sending a current i to a control element, particularly an electro-valve, which controls the clutch 21. The variation of the current i is shown with a dashed line in FIG. 4.

As the current i is increased, the clutch 21 starts to engage and the torque M increases. When the torque reaches a predefined value M_(T), increase of the current i is stopped and the current i is caused to decrease to a limited extent. This implies that the clutch 21 is partially disengaged, without however being completely opened. Thereafter, the current i is kept constant.

In this step, the clutch 21 is engaged up to approximately 25-35% of its capacity. In other words, pressure of a hydraulic fluid that is sent to the clutch 21 in order to engage the latter does not exceed 25-35% of the pressure that the hydraulic fluid has when the clutch 21 is fully engaged.

Excessive stress on the components of the transmission system 1 during the testing method can thus be avoided.

The torque M responds with a certain delay to the variations of the current i. Thus, when the current i starts decreasing, the torque M continues to increase for a while. Thereafter, also the torque M decreases slightly before stabilizing and remaining constant for a certain period.

As the torque M increases, also the pressure difference Δp between pressure of the hydraulic fluid in the high-pressure line 9 (or 10) and pressure of the hydraulic fluid in the low-pressure line 10 (or 9) increases.

The period in which the current i and the torque M increase and then slightly decrease before reaching a stable value is indicated by the letter C in FIG. 4.

This period can be considered as a transient phase in which a load is progressively applied to the transmission system 1. The load is due to the wheels 3 which, through the clutch 21, receive power from the engine 2 via the transmission system 1. The wheels 3 remain nevertheless still, since the vehicle is hand-braked.

Duration of period C is the result of a compromise between two opposite needs. On the one hand, period C should not be to long in order to avoid useless stress on the clutch 21. On the other hand, period C should not be too short in order to ensure that at its end stable conditions are reached.

As shown in FIG. 3, when the torque M is stabilized after engaging the clutch 21, a second measuring step 44 is provided in which several working parameters of the vehicle are measured.

In particular, the rotational speed n_(HST) of the output shaft 17 is measured by means of the sensor 18. The rotational speed n_(E) of the engine shaft 4 is also recorded by means of the control device which controls the engine 2. Also in this case, several values of the rotational speeds of the output shaft 17 and of the engine shaft 4 may be recorded and then processed to obtain an average value.

The rotational speeds of the output shaft 17 and of the engine shaft 4 measured in the second measuring step 44 (i.e., when a load is applied to the transmission system 1) are indicated respectively by n_(HST) and n_(E) to distinguish them from the values n_(HST0) and n_(E0) measured during the first measuring step 42, i.e. when no load was applied to the transmission system 1.

During the second measuring step 44, it is also possible to measure the value M_(D) of the torque M at the damper 13 by means of the detector 20.

Furthermore, the value M_(E) of the torque can also be recorded directly in the engine 2, by means of the control device which controls the engine 2.

The reason why M_(D) and M_(E) are recorded will be explained later.

The second measuring step 44 is carried out during the period indicated by the letter D in FIG. 4. This period can be considered as a stable phase in which a load is applied to the transmission system 1. As shown in FIG. 4, during the period D the torque M, as well as the current i, are substantially constant.

After the second measuring step 44 is concluded, the clutch 21 can be disengaged, so that both the torque M and the current i decrease.

The values that have been measured during the first measuring step 42 and during the second measuring step 44 can be processed as will be explained below.

In a calculating step 45, an actual value Q_(a) of a parameter Q which is indicative of the volumetric efficiency η_(v) of the hydrostatic unit 5 is calculated, particularly from the values of rotational speeds η_(HST0), n_(E0), n_(HST), n_(E).

In the embodiment that is being considered, the parameter Q is the ratio between the volumetric efficiency η_(v) of the hydrostatic unit 5 determined when a load is applied and the volumetric efficiency η_(v0) of the hydrostatic unit 5 when no load is present, i.e. when the clutch 21 is disengaged.

The general definition of volumetric efficiency η_(v) of a hydrostatic transmission is given below:

$\eta_{v} = \frac{n_{HM} \cdot V_{HM}}{n_{HP} \cdot V_{HP}}$

wherein:

-   -   n_(HM) is the rotational speed of the hydraulic motor of the         hydrostatic unit (output speed);     -   V_(HM) is the absorption volume of the hydraulic motor;     -   n_(HP) is the rotational speed of the hydraulic pump of the         hydrostatic unit (input speed);     -   V_(HP) is the displacement volume of the hydraulic pump.

In the exemplary embodiment discussed herein, the absorption volume of the hydraulic motor is constant, whereas the displacement volume of the hydraulic pump is variable and can be expressed as follows:

V _(HP) =α·V _(HPMAX)

where V_(HPMAX) is the maximum displacement volume that the hydraulic pump may have and α has been previously defined as the ratio between the current value of the swivel angle of the swash plate in the hydraulic pump and the maximum possible value of the swivel angle.

In the case at issue, the rotational speed n_(HM) of the hydraulic motor is the rotational speed of the output shaft 17 as measured by the sensor 18, i.e. the speed which has previously been called n_(HST) or n_(HST0), depending on whether a load was applied or not to the transmission system 1.

The rotational speed n_(HP) of the hydraulic pump is proportional to the rotational speed of the engine shaft 4 as recorded by the control device of the engine 2, through a coefficient which depends on the features of the transmission elements interposed between the hydraulic pump 6 and the engine 2. For a given geometry and layout of these transmission elements, the coefficient is constant and can be calculated. In other words, n_(HP) is proportional to the rotational speed that has previously been designated as n_(E) or n_(E0), depending on whether a load was applied or not to the transmission system 1.

The actual value Q_(a) of parameter Q, which is indicative of the volumetric efficiency of the hydrostatic transmission, can therefore be calculated as follows:

$Q_{a} = {\frac{\eta_{V}}{\eta_{V\; 0}} = \frac{n_{HST} \cdot n_{E\; 0}}{n_{HSTO} \cdot n_{E}}}$

The rotational speeds of the output shaft 17 have been determined respectively in the first measuring step 42 (n_(HST0)) and in the second measuring step 44 (n_(HST)) . Similarly, the rotational speeds of the engine shaft 4 have been determined respectively in the first measuring step 42 (n_(E0)) and in the second measuring step 44 (n_(E)).

Hence, at this point of the testing procedure, the actual value Q_(a) can be easily calculated.

The actual value Q_(a) can be calculated by using only measures of rotational speeds of components of the transmission system 1, particularly the output shaft 17 and the engine shaft 4. These measures can be taken in a simple way and by means of sensors or detectors that are not particularly complicated or expensive. Hence, calculating the actual value Q_(a) does not involve particular difficulties.

However, the actual value Q_(a) of the parameter Q, if taken alone, does not allow an operator to assess whether the hydrostatic unit 5 is working properly or not. This is because the parameter Q depends on the working conditions of the hydrostatic unit 5, and particularly on the pressure difference Δp between pressure of the hydraulic fluid in the high-pressure line 9 (or 10) and pressure of the hydraulic fluid in the low-pressure line 10 (or 9).

As shown in FIG. 5, the parameter Q generally decreases as the pressure difference Δp increases. Thus, a value of the parameter Q which is optimal for certain working conditions, and particularly for a given pressure difference Δp, could be problematic in different working conditions, i.e. for a different pressure difference Δp.

To be precise, the parameter Q may depend also on other factors which may change as the working conditions of the hydrostatic unit 5 change. These factors are, for example, the parameter α and the temperature of the hydraulic fluid flowing in the hydrostatic transmission. However, the parameter α has been set to a desired value in the initial setting step and is kept constant throughout the whole testing method. The testing method can be carried out in a temperature range in which the temperature of the hydraulic fluid has substantially no influence on the parameter Q.

Hence, the parameter Q may be considered as being influenced mainly by the pressure difference Δp.

The pressure difference Δp can be calculated once the features of the transmitting elements interposed between the engine 2 and the hydrostatic unit 5 are known, particularly once the features of the planetary gearing 22 are known.

It has been found that, in the configuration in which the testing method is carried out, the following relation exists between the pressure difference Δp and the torque M generated by the engine 2:

Δp=k·M

In other words, the pressure difference Δp is proportional to the torque M through a coefficient k.

This relation has been calculated mathematically and also experimentally confirmed by measuring both the pressure difference Δp and the torque M.

Therefore, starting from the value M_(D) of the torque as measured during the second measuring step 44 at the damper 13, it is possible to calculate a specific pressure difference Δp_(D) in the hydrostatic unit 5.

In order to assess whether the actual value Q_(a) of parameter Q corresponds to a properly working hydrostatic unit, at least one expected value of the parameter Q is determined in a determining step 46. The manner in which this can be done will be explained below.

One or more reference vehicles, of the same kind as the tested vehicle, have been previously tested in the same working conditions as those followed in the above disclosed testing procedure. The reference vehicles had transmission units which worked perfectly. The parameter Q has been calculated for the reference vehicles for a very high number of values of pressure difference Δp. In other words, a precise mapping of the parameter Q as a function of the pressure difference Δp (and hence as a function of the torque M, which is related to the pressure difference Δp as discussed above) has been carried out for the reference vehicles.

A table has thus been obtained which, for any selected value of the torque M, allows the corresponding value of the parameter Q to be determined in the case of a perfectly working hydrostatic unit. This table is stored in a control unit of the vehicle to be tested.

Starting from the value of the torque M_(D) which has been recorded in the second measuring step 44 by the detector at the damper 13, the control unit therefore determines, on the basis of the table derived from the reference vehicles, the expected value of the parameter Q, that will be indicated as Q_(ED) in the following. The expected value Q_(ED) is the value that the parameter Q should have if the hydrostatic unit 5 of the tested vehicle worked as the hydrostatic units of the reference vehicles.

The actual value Q_(a) of the parameter Q calculated for the vehicle under test and the expected value Q_(ED) can be further processed in order to assess whether the hydrostatic unit 5 of the vehicle under test is properly working.

For example, the actual value Q_(a) and the expected value Q_(ED) can be displayed so that an operator can evaluate how the hydrostatic unit 5 is working, in a displaying step 47. It is also possible to compare automatically the actual value Q_(a) with the expected value Q_(ED) and generate a message which informs the user about the status of the hydrostatic unit 5.

For example, if the actual value Q_(a) is 0.950 and the expected value Q_(ED) is 0.975, the operator can immediately see that the hydrostatic unit 5 has a volumetric efficiency which is roughly lower than the volumetric efficiency of the reference vehicles by 2.5%.

Thus, the hydrostatic unit 5 can be tested while it is still assembled on the vehicle and can be dismounted from the vehicle only if the testing procedure shows that the hydrostatic unit 5 has an abnormally low volumetric efficiency.

If, on the other hand, the testing procedure shows that the hydrostatic unit 5 has a volumetric efficiency—or more precisely, a value of a parameter which is indicative of the volumetric efficiency—sufficiently close to the reference vehicles, than the hydrostatic unit 5 does not need to be replaced and the fault is to be searched for in a different component.

It is also possible—either in addition to Q_(ED) or as an alternative—to determine the expected value of the parameter Q starting from the torque M_(E) recorded in the second measuring step 44 by the control device which controls the engine 2, i.e. upstream of the damper 13. This expected value will be indicated as Q_(EE) and is determined by inputting the torque M_(E) into the table predisposed on the basis of the reference vehicles.

By calculating the expected value Q_(ED) based on the engine torque M_(E), additional information on the transmission system 1 can be obtained.

For example, the following situation may occur:

Q_(a)=0.950 (actual value for the tested vehicle)

Q_(ED)=0.975 (expected value based on the torque at the damper 13)

Q_(EE)=0.975 (expected value based on the torque in the engine 2)

In this case, the operator can infer that the detector 20 is well calibrated, because the two expected values of the parameter indicative of volumetric efficiency do not differ substantially from one another.

The low actual value Q_(a) really corresponds to a low volumetric efficiency of the hydrostatic unit.

Another example can be the following:

Q_(a)=0.950 (actual value for the tested vehicle)

Q_(ED)=0.975 (expected value based on the torque at the damper 13)

Q_(EE)=0.950 (expected value based on the torque in the engine 2)

In this case, the detector 20 is not well calibrated, because the two expected values of the parameter indicative of volumetric efficiency are substantially different from one another.

The hydrostatic unit 5 is probably working properly, because the actual value Q_(a) corresponds to the expected value Q_(EE) based on the torque in the engine. However, the vehicle has probably a transmission with anomalous losses, due to the high expected value Q_(ED) based on the torque at the damper 13.

In any case, precious information can be obtained without disassembling the hydrostatic unit 5 unless this is truly necessary, which allows considerable savings in time and costs.

As mentioned before, a calibration procedure can be provided at the beginning of the testing method for calibrating the detector 20 located at the damper 13, so that the same mechanical losses are detected both for the tested vehicle and for one or more reference vehicles. These mechanical losses are due to the planetary gearing 22, the other toothed wheels, the pumps and more in general the mechanical components of the transmission system 1, which unavoidably has a mechanical efficiency lower than 1.

The detector 20 can be so configured as to read the value of a phase angle φ at the damper 13. The phase angle φ corresponds to the phase difference between the angular speed of the engine shaft 4 upstream of the damper 13 and the angular speed of the main shaft 12 downstream of the damper 13.

By measuring the phase angle φ, the torque at the damper 13 can be determined on the basis of a characteristic curve that, for each damper, gives the relationship between the phase angle φ and the torque.

The calibration step is carried out while the engine 2 is running, the clutch 21 is open and the vehicle is still. In this condition, the torque generated by the engine 2 is indicative of the mechanical losses of the transmission system 1, because no external device to be powered is attached to the power take-off of the vehicle.

In the calibration step, the behaviour of the tested damper 13 is compared to the behaviour of a damper of a reference vehicle.

For each kind of damper, a characteristic curve is known and is provided to the user by the damper manufacturer. This characteristic curve is shown by a continuous line in FIG. 6, which shows how the torque M_(D) varies as a function of the phase angle φ.

The dashed line of FIG. 6 shows, on the other hand, the characteristic curve of the tested damper, which is shifted with respect to the characteristic curve provided by the damper manufacturer by an offset Δφ_(av) that will be calculated as explained below.

For the purposes of the calibration step, it is assumed that each transmission system 1 has the same mechanical losses for a given temperature.

The behaviour of the damper to be calibrated is studied at different positions of the swash plate of the hydraulic pump 6. In other words, the behaviour of the damper to be calibrated is studied for different values of the parameter α, i.e. the ratio between the current value of the swivel angle of the hydraulic pump 6 and the maximum possible value of the swivel angle.

By varying the position of the swash plate, it is possible to obtain different losses in the hydrostatic unit 5. These losses originate different loads applied at the side of the damper 13 opposite the engine 2, i.e. different torques.

Typically, the losses obtained in the hydrostatic unit 5 by varying the position of the swash plate are of the order of 30-40 N·m. These losses are high enough to enable phase differences to be detected at the sides of the damper 13, without however having to engage the clutch 21. Hence, the clutch 21 is not stressed while the damper 13 is calibrated. Furthermore, the torque generated by the engine 2 in this step is very low. Substantially no wear is thus created on the transmission system 1.

In order to improve calibration accuracy, the behaviour of the damper 13 to be calibrated can be considered at a plurality of rotational speeds of the engine 2.

For example, five different values of the parameter α and three different values of the rotational speed of the engine shaft 4 can be taken into consideration. In this case, the behaviour of the damper 13 is studied at fifteen different calibration points P_(i), corresponding to the five values of the parameter α for each value of the rotational speed.

Each calibration point P_(i) defines a working condition of the transmission system 1.

For each calibration point P_(i), the phase angle φ_(i) of the damper 13 is recorded.

For each calibration point P_(i), it is known from the characteristic curve provided by the damper manufacturer which expected value φ_(ei) of the phase angle should be obtained.

To be precise, for each calibration point P_(i), a respective expected torque M_(EDi) should be theoretically obtained at the damper 13. The expected torque M_(EDi) can be derived from a table completed after studying one or more reference, well working vehicles. By inputting the expected torque M_(EDi) in the characteristic curve of the damper, as supplied by the damper manufacturer, the corresponding expected phase angle φ_(ei) can be found out.

However, for simplifying the procedure, it is possible to establish a direct relationship, for example by means of a suitable table or graph, between each working point P_(i) and the corresponding expected value φ_(ei) of the phase angle.

In any case, for each calibration point P_(i), it is possible to determine the difference or offset Δφ_(i) between the recorded phase angle φ_(i) of the damper to be calibrated and the expected phase angle φ_(ei).

An average offset Δφ_(av) can be calculated by averaging the single offsets Δφ_(i) determined for each calibration point P_(i). The average offset Δφ_(av) can be used to calibrate the detector 20 associated to the damper 13 by instructing the control unit to consider each measured value φ_(m) of the phase angle as if it were φ_(m)−Δφ_(av) and use the resulting phase angle to calculate the torque on the basis of the characteristic curve of the damper.

In this way, the torque detected at the damper 13 of the tested vehicle, for a given combination of rotational speed and α, will be the same as the torque detected by the damper of the reference vehicle, for that given combination of rotational speed and α. In other words, the losses of the transmission system 1 of the tested vehicle will be the same as the losses of the transmission system of the reference vehicle.

In an alternative embodiment, a temperature compensation can be adopted in the calibration procedure to take into consideration the effects of temperature of the hydraulic fluid, particularly oil, circulating in the hydrostatic unit 5. To this end, a temperature correction factor F_(TC) can be added to each recorded phase angle φ_(i) before determining the offset Δφ_(i).

In conclusion, the testing method shown in FIGS. 2 and 3 can be carried out fully automatically in order to determine, in few minutes, whether the hydrostatic unit 5 installed on the vehicle is properly working or not.

The calibration procedure can also be performed quickly and automatically, so as to increase precision of the test results.

However, the calibration procedure could also be carried out independently of the test method shown in FIGS. 2 and 3, each time the detector 20 associated to the damper 13 needs to be calibrated for whatsoever purpose.

In the calibration procedure which has been disclosed above, different hydraulic losses have been generated in the hydrostatic unit 5 by varying the swivel angle of the swash plate in the hydraulic pump 6.

In an embodiment which is not shown, the continuously variable transmission can comprise a hydraulic motor having a variable absorption volume. In this case, different hydraulic losses could be generated by varying the swivel angle of a swash plate of the hydraulic motor, in addition or as an alternative to varying the swivel angle in the hydraulic pump.

Instead of creating different torques at the damping assembly 31 by generating different hydraulic losses in the hydrostatic unit 5, it is also possible to generate different mechanical losses in a transmission system 1 that does not necessarily have to comprise a hydrostatic unit.

For example, FIG. 7 shows a transmission system 101 comprising a powershift transmission for transmitting power from an engine to the wheels. The transmission system 101 can be installed on a vehicle, particularly a working vehicle of any kind such as a tractor, an excavator, a crawler vehicle, a harvester or the like.

An engine which is not shown is arranged for rotating a driving shaft 104, which is connected to a transmission shaft 112 by a damping assembly 131. The damping assembly 131 comprises a damper 113 that can be associated to the transmission shaft 112. The damping assembly 131 further comprises a detector 120 for detecting the phase angle φ at the damper 113 and hence determining torque at the damper 113 by means of a phase angle-torque characteristic.

A first toothed wheel 140 is mounted coaxially with the transmission shaft 112. A first clutch 141 is interposed between the transmission shaft 112 and the first toothed wheel 140 so as to enable the first toothed wheel 140 to be selectively rotated by the transmission shaft 112.

A further transmission shaft 142 is aligned with the transmission shaft 112. The further transmission shaft 142 is located at an end of the transmission shaft 112 opposite the damping assembly 131. A second clutch 143 is interposed between the transmission shaft 112 and the further transmission shaft 142, so that the further transmission shaft 142 can be selectively rotated by the transmission shaft 112.

A plurality of toothed wheels are supported by the further transmission shaft 142, in a fixed manner relative to the further transmission shaft 142. In the embodiment of FIG. 7, three toothed wheels 144, 145, 146 are mounted on the further transmission shaft 142. In general, the toothed wheels supported by the further transmission shaft 142 have dimensions and geometry which differ from one toothed wheel to the other.

Both the transmission shaft 112 and the further transmission shaft 142 are hollow, so that a main shaft 147 extends internally of the transmission shaft 112 and the further transmission shaft 142. The main shaft 147 can be rotated by the driving shaft 104 via the damping assembly 131 independently of the transmission shaft 112 and the further transmission shaft 142.

A gear wheel 148 is keyed onto the main shaft 147 for driving one or more auxiliary pumps P shown only schematically in FIG. 7. The auxiliary pumps P are arranged for providing hydraulic power to a plurality of auxiliary devices such as, for example, a lift or one or more distributors.

A power take-off which is not shown is connected to an end of the main shaft 147 opposite the further end of the main shaft 147 connected to the damper 113.

The first toothed wheel 140 engages with a second toothed wheel 149 obtained in a component 150 supported by a shaft 151. On the component 150 a third toothed wheel 152 is also provided, which is arranged to move fixedly with the second toothed wheel 149. The third toothed wheel 152 engages with the toothed wheel 144 supported by the further transmission shaft 142.

A third clutch 153 is interposed between the shaft 151 and the component 150, so that the component 150 can either be rotated by the shaft 151 or disconnected from the shaft 151.

The shaft 151 additionally supports two further toothed wheels 154, 155, that are arranged to engage respectively with the toothed wheels 145, 146 mounted on the further transmission shaft 142.

A fourth clutch 156 is interposed between the shaft 151 and the further wheel 154, so that the further wheel 154 can be either coupled to, or disconnected from, the shaft 151. Similarly, a fifth clutch 157 is interposed between the shaft 151 and the further wheel 155, in order to enable connection or disconnection between the shaft 151 and the further wheel 155.

The shaft 151 is connected to a gearing assembly 158 comprising a plurality of toothed wheels and a plurality of clutches 159.

A driven shaft 160 is provided at the output of the gearing assembly 158. The driven shaft 160 is connected, through a plurality of known devices that are not shown, to the wheels of the vehicle in which the transmission system 101 is included.

By acting on the clutches 159 of the gearing assembly 158, different travelling modes can be selected for the driven shaft 160, for example slow, medium, fast and reverse.

By acting on the clutches 141, 143, 153, 156, 157, different transmission ratios can be selected as an input for the gearing assembly 158. For example, the first clutch 141 and the third clutch 153 could be closed, while the second clutch 143, the fourth clutch 156 and the fifth clutch 157 are left open. In this case, torque goes to the gearing assembly 158 through the first toothed wheel 140 and the second toothed wheel 149.

If, on the other hand, the first clutch 141 and the fourth clutch 156 are closed, while the second clutch 143, the third clutch 153 and the fifth clutch 157 are left open, torque goes to the gearing assembly 158 through a more complicated path. This path passes through the first toothed wheel 140, the second toothed wheel 149, the third toothed wheel 152, the wheels 144, 145 and finally the further toothed wheel 154.

Many other combinations of toothed wheels 140, 149, 152, 144, 145, 146, 154, 155 and of clutches 141, 143, 153, 156, 157 can be chosen in order to have different transmission ratios and torque paths between the damping assembly 131 and the gearing assembly 158. Each of these combinations creates different losses in the transmission system 101, due for example to different frictions arising between the involved clutches and toothed wheels. Consequently, different torques are generated at the side of the damping assembly 131 facing the transmission system 101.

A plurality of working conditions or calibration points P_(i) are thus obtained, for each of which an actual value of the phase angle φ_(i) can be measured by the detector 120. The number of calibration points P_(i) can be increased by varying the rotational speed of the engine and hence of the driving shaft 104.

The values of phase angle φ_(i) measured at the so obtained calibration points P_(i) can be subsequently processed in order to calibrate the detector 120, in the same manner as was previously disclosed for the embodiment in which different calibration points P_(i) were obtained by varying the swivel angle.

By selecting an appropriate number of calibration points P_(i), the detector 120 can be accurately calibrated even in a transmission system which does not comprise a hydrostatic unit.

Furthermore, during the calibration procedure the clutches 159 are left open or disengaged, so that the wheels are disconnected from the engine and the vehicle is still. Hence, during the calibration procedure no wear is generated on the clutches 159.

Since the wheels are disconnected, the torque that is generated in the transmission system 1 during the calibration procedure is limited, which avoids excessive stress on components of the transmission system 1.

In an embodiment which is not shown, suitable synchronizers could be used in place of the clutches 159, in which case the wheels would be disconnected from the engine during the calibration procedure by disengaging the synchronizers. 

1. A method for calibrating a detector of a damping assembly in a vehicle, the damping assembly comprising a damper for connecting a transmission shaft of a transmission system of the vehicle to a driving shaft, comprising the step of applying different torques at the damping assembly by generating losses at different working conditions of the transmission system, so that said damper can be compared to a damper of a reference vehicle for each of said different working conditions, wherein the transmission system is disconnected from an axle of the vehicle during the step of applying different torques, the axle supporting wheels or crawler sprockets of the vehicle, wherein the transmission system is disconnected from the axle by opening at least one clutch and/or at least one synchronizer interposed between the axle and the transmission system, and wherein the losses at different working conditions are generated by selecting different positions of a swash plate of a variable displacement hydraulic pump and/or a variable absorption hydraulic motor included in the transmission system, each position corresponding to a selected swivel angle of the swash plate.
 2. (canceled)
 3. (canceled)
 4. (canceled)
 5. A method according to claim 1, wherein the losses at different working conditions are generated by selecting different positions of a swash plate of a variable displacement hydraulic pump and/or a variable absorption hydraulic motor included in the transmission system, each position corresponding to a selected swivel angle of the swash plate.
 6. A method according to claim 1, wherein the losses at different working conditions are generated by selecting different torque paths for transmitting torque between a plurality of toothed wheels of the transmission system.
 7. A method according to claim 5, wherein the different torque paths are selected by engaging or disengaging predefined clutches interposed between said toothed wheels.
 8. A method according to claim 1, wherein the losses at different working conditions are further generated by selecting different rotational speeds of the driving shaft.
 9. A method according to claim 1, wherein said damper is compared to the damper of the reference vehicle by determining the difference (φ_(i)) between a phase angle value (φ_(i)) measured by the detector and a phase angle value (φ_(ei)) of the damper of the reference vehicle, for each of said different working conditions.
 10. A method according to claim 8, wherein the phase angle values (φ_(ei)) of the damper of the reference vehicle are determined by means of a characteristic curve establishing a relationship between torque and phase angle.
 11. A method according to claim 9, wherein each phase angle value (φ_(ei)) of the damper of the reference vehicle is determined by inputting in the characteristic curve a torque recorded for the reference vehicle in the corresponding working condition.
 12. A method according to 10, and further comprising the step of averaging the differences (φ_(i)) between the phase angle values (φ_(i)) of said damper and the phase angle values (φ_(ei)) of the damper of the reference vehicle, so as to obtain an average difference (φ_(AV)).
 13. A method according to claim 11, and further comprising the step of instructing a control unit controlling said damper to offset each subsequently measured value (φ_(m)) of the phase angle by an amount equal to said average difference (φ_(AV)).
 14. A method according to claim 12, and further comprising a temperature calibration step in which a temperature correction factor (F_(TC)) is added to each phase angle value (φ_(i)) of said damper, the temperature correction factor (F_(TC)) taking into consideration the effects of temperature of a hydraulic fluid circulating in the transmission system.
 15. A method according to claim 13, wherein the phase angle is the phase difference between rotational speed of the transmission shaft and rotational speed of the driving shaft.
 16. A method according to claim 1, wherein the method is accomplished via a computer program product, which program is operative to cause a processor to perform the method. 